1. Field of the Invention
The present invention relates to an externally pressurized gas bearing incorporated, for example, into a precision processing machine or inspection device to support the main shaft in a non-contact state, and it also relates to a spindle device using this.
2. Brief Description of the Prior Art
An externally pressurized gas bearings, which supports the main shaft in a non-contact state with respect to the bearing surface, has a high degree of rotation accuracy, and is used in the work spindle device or tool spindle device of a precision processing machine or precision inspection device. Prior art examples of such externally pressurized gas bearing and a spindle device using this are shown in FIG. 9.
This externally pressurized gas bearing comprises bearing sleeves 4, 5 and 6 fixed in a housing 3 by a suitable means, a main shaft 1 that is radially supported in a non-contact state through a fine bearing clearance by journal bearing sections 7 and 8 formed by the two bearing sleeves 4 and 5, and that is axially supported in a non-contact state through a fine bearing clearance by a pair of thrust bearing sections 9 and 10 formed by holding a thrust plate 2 integral with the main shaft 1 between two bearing sleeves 5 and 6 from opposite surfaces. A spindle device having this externally pressurized gas bearing is provided with a driving source 22 coaxial with the main shaft 1, said driving source 22 having a motor rotor 11 integrally attached to the main shaft 1 and adapted to rotate the main shaft 1 by a driving force produced by an electromagnetic force acting between itself and a motor stator 12.
The bearing sleeves 4 and 5 constituting the journal bearing sections 7 and 8 are respectively provided with two gas feed rows 23 each having a plurality of circumferentially equispaced fine gas feed holes 13 that open to the bearing surface opposed to the main shaft 1. Further, the outer surface of the main shaft 1 is formed with circumferential grooves 18 at positions opposed to two gas feed rows 23 formed in the bearing sleeves 4 and 5.
The bearing sleeves 5 and 6 constituting the thrust bearing sections 9 and 10 are respectively provided with a single circumferential gas feed row 24 having a plurality of circumferentially equispaced fine gas feed holes 14 that open to the bearing surface opposed to the thrust plate 2. In some cases, in order to increase the static stiffness of the thrust bearing sections 9 and 10, a circumferential groove is formed that connects gas feed holes 14 in the gas feed row 24 formed in the bearing sleeves 5 and 6.
In this externally pressurized gas bearing, when compressed gas is fed from a bearing gas feed port 15, it flows into the bearing clearances in the journal bearing sections 7 and 8 and thrust bearing sections 9 and 10 from the gas feed holes 13 and 14 in the gas feed rows 23 and 24 via a gas feed passageway 16 formed in the housing 3, and a load carrying force of bearing is produced that balances with the self-weight of the main shaft 1 and external load by the pressure of the compressed gas in the bearing clearances. With these journal bearing sections 7 and 8 and thrust bearing sections 9 and 10, the main shaft 1 is driven for rotation while being supported in a non-contact state, whereby a highly accurate rotary motion is realized. In addition, the gas flowing out of the journal bearing sections 7 and 8 and thrust bearing sections 9 and 10 is discharged to the outside of the housing 3 directly from the bearing ends or through exhaust passageways 17.
A seal sleeve 19 is disposed between the bearing sleeves 4 and 5. This seal sleeve 19 is formed in its inner and outer surfaces with suction holes 25 extending between the inner and outer surfaces and a circumferential groove 26 communicating with the suction holes 25. An suction passageway 21 formed in the main shaft 1 and a suction passageway 20 formed in the housing 3 communicate with each other through the suction holes 25 and circumferential groove 26. The opposite sides of the circumferential groove 26 formed in the inner surface of the seal sleeve 19 are opposed to the outer diameter surface of the main shaft 1 through the same fine seal clearance as the bearing clearance in the journal bearing sections 7 and 8, thus presenting a non-contact seal construction.
In using this spindle device, a vacuum chuck (not shown) or the like is attached to the front end of the main shaft 1, and is used by evacuating the device through the exhaust passageway 21 in the main shaft 1 via the seal sleeve 19 by an external vacuum pump (not shown) connected to the exhaust passageway 20.
In this connection, the radial runout accuracy of the main shaft 1 is influenced mainly by the characteristics of the journal bearing sections 7 and 8. Further, since decreasing the bearing clearance is effective in increasing the stiffness and damping coefficient of the externally pressurized gas bearing, it is common practice to set the bearing clearance at as small a value as possible within the range that is permitted by other factors and to determine the size and number of the gas feed holes 13 and the axial position of the gas feed rows 23 so that the bearing stiffness may be greatest with this clearance. In this case, if the bearing clearance is decreased, the resistance of the bearing clearance to the flow of the compressed gas increases, and it becomes necessary to correspondingly increase the resistance of the gas feed holes 13, which means that the diameter of the gas feed holes 13 should be minimized and the number of gas feed holes should be decreased.
Further, no special consideration has heretofore been paid to the circumferential position of the gas feed holes 13 in the gas feed rows 23 in the journal bearing sections 7 and 8. Thus, the gas feed holes 13 in the two gas feed rows 23 in the journal bearing sections 7 and 8 are disposed in the same phase (the same circumferential position), and the phase relationship of the gas feed holes 13 in the journal bearing sections 7 and 8 are not specially prescribed in most cases.
In recent years, as semiconductors and information mediums become increasingly densified and increasingly microscopic in structure, it has been desired to further improve the runout accuracy of the main shaft for the externally pressurized gas bearing in consideration of the fact that particularly the radial runout of the main shaft has a great influence on the test and processing accuracies. Concerning the main shaft runout accuracy of this externally pressurized gas bearing, it has been theoretically shown, when an externally pressurized gas bearing of inherent restrictor type having a single gas feed row without a circumferential groove is used as a subject of analysis (bearing model), that radial runout is produced during rotation of the main shaft by the generation of exciting force of particular frequency caused by the interaction between the number of gas feed holes and the main shaft shape (roundness error), as disclosed in “Fundamental Study on Rotation Accuracy Characteristics of Externally Pressurized Gas Journal Bearings (second report, Shaft Rotation Accuracy Characteristics)” (Transaction of the JSME (Series C), Vol. 58, No. 548 (1992–4), pp. 1177–1183).
This paper points out that in the case where the number of gas feed holes constituting the gas feed row is k, there is a possibility that radial runout of the main shaft having a frequency equal to n k±1 (where n=1, 2, . . . ) times the rotation speed will occur due to the interaction with the roundness error of the main shaft. Concerning the influences of rotation speed of the main shaft, it is also shown that if the rotation speed increases until the frequency of runout of the main shaft exceeds the resonance point, the amplitude of runout decreases. Theoretically, therefore, it is presumed that the externally pressurized gas bearing that is provided with the two gas feed rows 23 and circumferential groove 18 as shown in FIG. 9 will induce the same phenomenon.
In order to reduce the runout of the main shaft due to interaction between the number of gas feed holes and the main shaft shape (roundness error), the provision of the circumferential groove 18 in the outer surface of the main shaft 1, as shown in FIG. 9, combined with the increased clearance in the gas feed hole outlet port and the circumferential leveling of pressure in the bearing clearance, can be expected to produce some effect. In order to produce a further effect in consideration of the fact that it is desired to further improve the runout accuracy of the main shaft 1 for the externally pressurized gas bearing as semiconductors and information mediums become increasingly densified and increasingly microscopic in structure, if the number k of gas feed holes in one gas feed row is increased until k−1 times of the specified rotation speed exceeds the resonant speed of the spindle device, then it is possible to reduce the amplitude of the runout. However, increasing the number of gas feed holes requires increasing the bearing clearance, resulting in incurring decreases in the bearing stiffness and in damping coefficient and an increase in the consumption of compressed gas. Particularly, decreases in stiffness and in damping coefficient present the problem of increasing the runout of the main shaft due to disturbance vibration.